Vibration reducing head suspension mechanism for a magnetic disc unit

ABSTRACT

There is provided a head suspension mechanism capable of more effectively restraining sway mode and second-order torsion mode vibrations of the head suspension mechanism caused by an air flow. A flange portion having a free end portion  23   c  on the base part side is formed at both edges close to the tip end of a suspension main frame  23 , and at the same time, an elastic material  23   d  is provided in the free end portion  23   c . Also, a sway mode dynamic vibration absorber is formed by the length of the flange portion and the damping effect of the elastic material  23   d.

This application is a continuation of co-pending application Ser. No.11/151,525, filed on Jun. 14, 2005. The entire contents are herebyincorporated by reference.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a head suspension mechanism which isused for a magnetic disc unit such as a hard disc mounted on a personalcomputer and elastically supports a head slider to perform recording andregenerating to a magnetic disc. More particularly, it relates to a headsuspension mechanism with a dynamic vibration absorber, which is capableof restraining vibrations caused by an air flow generating due to thehigh-speed rotation of the magnetic disc, especially sway modevibrations and second-order torsion mode vibrations that result in apositional shift between a head and a recording track.

2. Description of the Related Art

A personal computer and the like have conventionally used a magneticdisc unit as a recording/regenerating medium for information.

FIG. 12 is a plan view showing one example of a general magnetic discunit. As shown in FIG. 12, the magnetic disc unit D is configured sothat a magnetic disc body d capable of rotating at a high speed is heldin a housing B consisting of a lid-shaped mating structure (only one isshown), and a head arm HA, which is capable of being rotated in asubstantially radial direction (refer to an arrow mark) of the magneticdisc body d by the drive of a voice coil motor (VCM) M, is provided soas to face to the magnetic disc body d. Also, in a free end part of thehead arm HA, there is provided a head slider suspension HS that holds ahead slider (not shown) for performing writing of data to the magneticdisc body d and reading of the written data.

In recent years, the capacity and speed of the magnetic disc body d haveincreased. Accordingly, the head slider suspension HS is required tohave vibration damping properties so as to prevent the occurrence ofvibrations of head caused by an air flow (turbulent flow) generatingwhen the magnetic disc body d rotates.

FIG. 13 is a perspective view of a head suspension, showing one exampleof a conventional head suspension mechanism. This head suspension 1includes a fixed plate 2 stakingly connected to the aforementioned headarm HA, a suspension main frame 3 consisting of a thin metal sheet fixedto the fixed plate 2, a flexure 4 fixed at the tip end of the suspensionmain frame 3, and a head slider 5 fixed on the bottom surface at the tipend of the flexure 4.

In a base part of the suspension main frame 3, a leg portion 3 a isformed so as to provide flexural elasticity by blanking a central partthereof. Also, at both edges of the suspension main frame 3, flangeportions 3 b erecting along the edges are formed.

In the flexure 4, a substantially U-shaped slit 4 a is formed so as tosurround a central part thereof, so that by elastically fixing the headslider 5, a load from the suspension main frame 3 is applied to acentral part of the back surface (fixed surface) of the head slider 5.

Since the configuration is as described above, the head suspensionmechanism supports the head slider 5 for performingrecording/regenerating on the surface of the magnetic disc body drigidly in the in-plane direction (horizontal direction=XY direction)and flexibly in the out-of-plane direction (vertical direction=Zdirection), and also can perform a function of giving a fixed load forceto the head slider 5.

Also, since the head suspension mechanism is moved in the in-planedirection at a high speed to move the head slider 5 to a recording trackof the magnetic disc body d at a high speed, it is important that thesuspension mechanism be as light in weight as possible and moreover benot subjected to vibrations as an elastic body.

For this reason, as shown in FIG. 13, the suspension main frame 3 isformed into a tapered triangular shape by using a stainless steelmaterial with a thickness of several tens of micrometers, and the flangeportions 3 b erecting vertically are formed at both edges of thesuspension main frame 3, by which the suspension main frame 3 itself isconfigured so as to be light in weight and have high flexural rigidity.The flexible flexural elasticity function of the head suspensionmechanism is provided in the leg portion 3 a for fixing the suspensionmain frame 3 to the fixed plate 2.

Since the head slider 5 is fixed to the flexure 4, the head slider 5 issupported flexibly not only in the out-of-plane direction but also thepitching direction and the rolling direction so as to follow thevibrations and swell of the magnetic disc surface by means of air filmrigidity.

FIG. 14 is a perspective view of a head suspension, showing anotherexample of the conventional head suspension mechanism. In this headsuspension, the thickness of the suspension main frame 13 is increasedas a whole, by which the flange portions 3 b of the suspension mainframe 3 are eliminated. In FIG. 14, the same reference characters areapplied to elements that are substantially the same as those in FIG. 13,and the explanation thereof is omitted.

In recent years, the density of track recorded on the magnetic discsurface has been becoming high. Accordingly, an air flow around thesuspension, which is generated by the rotation of the magnetic disc bodyd, induces minute elastic vibrations on the order of nanometers, whichbecomes a main cause for a track positioning error of the head slider 5.

Such an elastic vibration has many components having a frequency higherthan a servo band, and hence cannot be restrained by control, so that anexcitation vibration itself must be reduced. Therefore, it is consideredthat as shown in FIG. 15, a restraining multilayer visco-elastic plate 6is affixed to the upper surface of the suspension main frame 3 toprovide damping.

This restraining multilayer visco-elastic plate 6 is formed into amultilayer structure in which the upper layer is a restraining metalplate 6 a and the lower layer is a visco-elastic material 6 b, and has aconstruction such that the visco-elastic material 6 b is held betweentwo metal plates of the suspension main frame 3 and the restrainingmetal plate 6 a.

The restraining multilayer visco-elastic plate 6 produces a dampingforce by means of relative motion between the visco-elastic material 6 band the upper and lower metal plates, so that an effect of dampingflexural vibrations of the suspension mechanism is great. However, therestraining multilayer visco-elastic plate 6 has a problem in that aneffect of damping torsional vibrations and sway mode vibrations in whichthe tip end of suspension vibrates swingingly in a plane is little.

The sway mode vibration is a vibration of a mode in which the headslider 5 is vibrated in the track position shift direction (Y directionin FIG. 15), and is responsible for a shift of track of the head slider5. Therefore, it is desirable to effectively damp the sway modevibrations excited by an air flow.

The torsional vibrations can also be adjusted so as not to affect thetrack position sift and float clearance variation due to the first-ordermode. However, the second-order torsion mode vibrations often produce atrack shift.

Thereupon, it is considered that the sway mode and second-order torsionmode vibrations of the head suspension mechanism caused by an air floware prevented (refer to U.S. Pat. No. 5,943,191).

FIG. 16 shows a head suspension showing an idea of applying a techniquesimilar to the dynamic vibration absorber to the head suspensionmechanism.

In the typical example shown in FIG. 16, a fixture 14 is provided with amustache-shaped damping plate 14 a, and a frictional force due torelative vibrations of a contact portion between the damping plate 14 aand a contact plate 15 caused when the flexure 14 is going to vibrate isutilized.

Such measures are effective when the flexure 14 resonates greatly.However, the measures have a drawback in that no effect is achieved inthe case of minute vibration in which the contact portion between thedamping plate 14 a and the contact plate 15 does not shift relatively.

SUMMARY OF THE INVENTION

The present invention has been made to solve the above problems, andaccordingly an object thereof is to provide a head suspension mechanismcapable of restraining, more effectively, sway mode and second-ordertorsion mode vibrations of the head suspension mechanism caused by anair flow.

To achieve the above object, an invention of a first aspect provides ahead suspension mechanism which supports a head slider rigidly in thein-plane direction and softly in the out-of-plane direction, and isprovided with a suspension main frame formed of an elastic cantileverthin sheet, which gives a load force to the slider, wherein a flangeportion erecting substantially vertically is formed at both edges near asupport portion of the head slider of the suspension main frame; aportion in which the flange portion and the suspension main frame arecontinuous is limited to the tip end portion of the suspension mainframe located near the head slider and the remaining portion of theflange portion is made a free end portion; and a sway mode dynamicvibration absorber is formed by the length of the flange portion and thedamping effect of an elastic material provided in the free end portion.

An invention of a second aspect provides a head suspension mechanismwhich supports a head slider rigidly in the in-plane direction andsoftly in the out-of-plane direction, and is provided with a suspensionmain frame formed of an elastic cantilever thin sheet, which gives aload force to the slider, wherein a pair of cut-and-raise shaped dualtongue-shaped elastic plates are provided at positions symmetrical withrespect to a center axis in almost the same direction as the extensiondirection of the suspension main frame near an antinode of second-ordertorsion mode vibration on the base part side of the suspension mainframe; and a second-order torsion mode dynamic vibration absorber isformed by the length of the dual tongue-shaped elastic plate and thedamping effect of a damping material provided on the surface thereof.

An invention of a third aspect provides a head suspension mechanismwhich supports a head slider rigidly in the in-plane direction andsoftly in the out-of-plane direction, and is provided with a suspensionmain frame formed of an elastic cantilever thin sheet, which gives aload force to the slider, wherein a flange portion erectingsubstantially vertically is formed at both edges near a support portionof the head slider of the suspension main frame; a portion in which theflange portion and the suspension main frame are continuous is limitedto the tip end portion of the suspension main frame located near thehead slider and the remaining portion of the flange portion is made afree end portion; and a sway mode dynamic vibration absorber is formedby the length of the flange portion and the damping effect of an elasticmaterial provided in the free end portion, and a pair of cut-and-raiseshaped dual tongue-shaped elastic plates are provided at positionssymmetrical with respect to a center axis in almost the same directionas the extension direction of the suspension main frame near an antinodeof second-order torsion mode vibration on the base part side of thesuspension main frame; and a second-order torsion mode dynamic vibrationabsorber is formed by the length of the dual tongue-shaped elastic plateand the damping effect of a damping material provided on the surfacethereof.

An invention of a fourth aspect provides a head suspension mechanism inwhich a visco-elastic material is provided between the free end portionand the suspension main frame.

An invention of a fifth aspect provides a head suspension mechanism inwhich a visco-elastic material is provided between the dualtongue-shaped elastic plate and the suspension main frame.

An invention of a sixth aspect provides a head suspension mechanism inwhich a vibration damping plate material consisting of a visco-elasticmaterial layer and a metal sheet material layer is provided on thesurface of the suspension main frame.

According to the head suspension mechanism in accordance with thepresent invention, the flange portion having the free end portion on thebase part side is formed at both edges close to the tip end of thesuspension main frame, and at the same time, the elastic material isprovided in the free end portion, and also, the sway mode dynamicvibration absorber is formed by the length of the flange portion and thedamping effect of the elastic material. Therefore, sway mode vibrationsof the head suspension mechanism caused by an air flow can be restrainedmore effectively.

Also, according to the head suspension mechanism in accordance with thepresent invention, the paired cut-and-raise shaped dual tongue-shapedelastic plates symmetrical with respect to the center axis are providedin locations close to the base part of the suspension main frame, and atthe same time, the damping material is provided in the free end portionof the dual tongue-shaped elastic plate, and also, the second-ordertorsion mode dynamic vibration absorber is formed by the length of thedual tongue-shaped elastic plate and the damping effect of the dampingmaterial. Therefore, second-order torsion mode vibrations of the headsuspension mechanism caused by an air flow can be restrained moreeffectively.

Further, the flange portion having the free end portion on the base partside is formed at both edges close to the tip end of the suspension mainframe, and at the same time, the elastic material is provided in thefree end portion, and also, the sway mode dynamic vibration absorber isformed by the length of the flange portion and the damping effect of theelastic material; and the paired cut-and-raise shaped dual tongue-shapedelastic plates symmetrical with respect to the center axis are providedin locations close to the base part of the suspension main frame, and atthe same time, the damping material is provided in the free end portionof the dual tongue-shaped elastic plate, and also, the second-ordertorsion mode dynamic vibration absorber is formed by the length of thedual tongue-shaped elastic plate and the damping effect of the dampingmaterial. Therefore, sway mode vibrations and second-order torsion modevibrations of the head suspension mechanism caused by an air flow can berestrained more effectively.

Also, the visco-elastic material is provided between the free endportion and the suspension main frame or between the dual tongue-shapedelastic plate and the suspension main frame. Therefore, an effect ofrestraining the sway mode vibrations and second-order torsion modevibrations of the head suspension mechanism caused by an air flow canfurther be increased.

Also, the vibration damping plate material consisting of thevisco-elastic material layer and the metal sheet material layer isprovided on the surface of the suspension main frame. Therefore, aneffect of restraining the sway mode vibrations and second-order torsionmode vibrations of the head suspension mechanism caused by an air flowcan further be increased.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a view showing an embodiment of a head suspension mechanismwith a dynamic vibration absorber for restraining sway mode vibrationsin accordance with the present invention, FIG. 1(A) being a perspectiveview of a head suspension, FIG. 1(B) being an enlarged side view showingone example of a principal portion, and FIG. 1(C) being an enlarged sideview showing another example of the principal portion;

FIG. 2 is an enlarged view showing a construction and configuration ofan L-shaped dynamic vibration absorber;

FIG. 3 is an enlarged view showing an operational function of anL-shaped dynamic vibration absorber;

FIG. 4 is a perspective view of a head suspension, showing anotherembodiment of a head suspension mechanism with a dynamic vibrationabsorber for restraining sway mode vibrations in accordance with thepresent invention;

FIG. 5 is a perspective view of a head suspension, showing an embodimentof a head suspension mechanism with a dynamic vibration absorber forrestraining second-order torsion mode vibrations in accordance with thepresent invention;

FIG. 6 is a perspective view of a head suspension, showing an embodimentof a head suspension mechanism with a dynamic vibration absorber, inwhich a plurality of dynamic vibration absorbers are installed torestrain sway mode and second-order torsion mode vibrations at the sametime;

FIG. 7 is an enlarged view of a slider;

FIG. 8 is a graph of an amplitude characteristic of a compliancefrequency response function in the Y direction (FIG. 8(A)) and in the Zdirection (FIG. 8(B)) at a head position of a head suspension mechanismwith a dynamic vibration absorber for damping sway mode resonance;

FIG. 9 is a graph of an amplitude characteristic of a compliancefrequency response function in the Y direction (FIG. 9(A)) and in the Zdirection (FIG. 9(B)) at a head position of a head suspension mechanismwith a dynamic vibration absorber for damping second-order torsion moderesonance;

FIG. 10 is a graph of an amplitude characteristic of a compliancefrequency response function in the Y direction (FIG. 10(A)) and in the Zdirection (FIG. 10(B)) at a head position of a head suspension mechanismwith a dynamic vibration absorber, in which two types of dynamicvibration absorbers are installed to damp sway mode and second-ordertorsion mode resonance;

FIG. 11 is a perspective view of a head suspension, showing an examplein which a restraining multilayer visco-elastic material is affixed tothe embodiment of the present invention shown in FIG. 6;

FIG. 12 is a plan view of a hard disc unit;

FIG. 13 is a perspective view of a head suspension, showing an exampleof a conventional head suspension mechanism;

FIG. 14 is a perspective view of a head suspension, showing anotherexample of a conventional head suspension mechanism;

FIG. 15 is a perspective view of a head suspension, showing an exampleof a conventional head suspension mechanism having a restrainingvisco-elastic damping plate material; and

FIG. 16 is a perspective view of a head suspension, showing an exampleof a conventional head suspension mechanism having a dynamic vibrationabsorber.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

A head suspension mechanism in accordance with the present inventionwill now be described with reference to the accompanying drawings.

First Embodiment

FIG. 1 is a perspective view of a head suspension, showing a firstembodiment of a head suspension mechanism in accordance with the presentinvention. In FIG. 1 and the following figures, the same referencecharacters are applied to elements that are substantially the same asthose in the related art, and the explanation thereof is omitted.

A head suspension 21 in accordance with the present invention has asuspension main frame 23 formed of a thin metal sheet.

In a base part of the suspension main frame 23, a leg portion 23 a isformed so as to provide flexural elasticity by blanking a central partthereof. Also, at both edges of the suspension main frame 23, flangeportions 23 b erecting along the edges are formed. Further, a free endportion 23 c that is made free toward the base part side by cuttingsubstantially into an L shape is formed in a location close to the tipend of the flange portion 23 b. The free end portion 23 c may have aconfiguration such that a rising portion ranging from the suspensionmain frame 23 to the flange portion 23 b is left as shown in FIG. 1(B),or may have a configuration such as to be formed into a flange shapeindependent of the flange portion 23 b without the rising portionranging from the suspension main frame 23 to the flange portion 23 bbeing left as shown in FIG. 1(C). A length L1 close to the tip end ofthe flange portion 23 b including the free end portion 23 c and a lengthL2 of the free end portion 23 c are set according to sway modevibrations generated by a rotational force etc. of a magnetic disc bodyd. At this time, the length L2 of the free end portion 23 c is set morestrictly than the length L1 close to the tip end of the flange portion23 b. On the other hand, on the outer surface side (or the inner surfaceside) of the free end portion 23 c, a damping material 23 d is providedas shown in FIG. 2.

Thereupon, the free end portion 23 c formed integrally of the samematerial as that of the suspension main frame 23 and the dampingmaterial 23 d affixed onto one-side surface of the free end portion 23 cconstitute a dynamic vibration absorber. The affixing region of thedamping material 23 d is determined so as to have a value such that thedamping effect is optimal, and the damping material 23 d can be affixedonto the inner surface of the free end portion 23 c as necessary.

FIG. 3 is a view more detailedly showing the vibration mode andconstruction when the free end portion 23 c operates as a dynamicvibration absorber. In FIG. 3, the illustration of the damping material23 d shown in FIG. 2 is omitted. The free end portion 23 c consists of abeam part a and a connecting part b. The beam part a has a property ofbeing subjected to flexural vibrations in the Y direction (the in-planedirection of the suspension main frame 23 or the track shift direction).Also, the connecting part b has a function of enhancing a spring effectof flexural vibration of the free end portion 23 c. FIG. 3 shows theflexural vibration form of the beam part a when it functions as adynamic vibration absorber and the deformation of the tip end of theconnecting part b corresponding to the flexural vibration form. Thesuspension side of the connecting part b is integrally connected to thesuspension main frame 23, so that it vibrates in the Y directionintegrally with the suspension main frame 23 when the head suspensionmechanism is subjected to sway mode vibrations in the Y direction.Therefore, as shown in FIG. 3, the displacement of the connectingportion between the connecting part b and the beam part a is a bendingdisplacement δ in the Y direction, and a torsional angle α in thedirection in which the beam part b is bent also contributes to thebending deformation.

Regarding the bending frequency of the free end portion 23 c, the lengthL2 of the beam part a is especially important, and it is important tomake design so that the first-order bending frequency approximatelycoincides with the sway mode frequency of the head suspension mechanismin accord with the theory of dynamic vibration absorber. Specifically,when viewing a frequency response characteristic of amplitude in the Ydirection of the tip end of the head suspension mechanism bydisplacement vibrating the fixed part of the head suspension mechanismor by force vibrating the suspension main frame 23, the length L2 of thebeam part a is determined so that two fixed points P and Q appearingnear the sway mode resonance frequency have almost the same height. Thisis referred to as “the length L2 of the beam part a is determined so asto meet the coincidence condition in accord with the theory of dynamicvibration absorber”.

After the length L2 of the beam part a is determined so as to meet theaforementioned coincidence condition, the free end portion 23 c requiresproper damping. To provide damping, various methods can be considered. Afirst method is to bond the damping material 23 d on one surface or bothsurfaces of the free end portion 23 c as shown in FIG. 2. As describedlater, the damping effect may be about 10⁻⁶ in terms of tan δ ofmaterial, and therefore a visco-elastic material has only to be bondedor applied.

A second method for providing damping, as shown in FIG. 4, is to fill avisco-elastic material 24 in part or in whole of a separation clearancebetween the beam part a and the suspension main frame 23 to provide adamping effect. The filling region differs depending on the effect ofthe damping material 23 d and the sway mode frequency, and thus theoptimum amount must be determined experimentally according to the designconditions of head suspension mechanism and the properties of thevisco-elastic material 24.

After the length L2 of the beam part a is determined so as to meet theaforementioned coincidence condition, the damping allowed to act on thebeam part a has an optimum value. If the damping is too large, aresonance peak appears in a central part of two resonance frequencies.Inversely, if the damping is too small, the peaks of two resonancefrequencies increase. Therefore, if the damping is determined so thatthe peaks of two resonance frequencies are in close proximity to thefixed points P and Q, the peaks near the two resonance frequencies withrespect to the disturbance can be made at a minimum at the same time.Such an optimum condition of damping effect should be determinedexperimentally because the optimum condition depends on the method forproviding damping and the damping material used. Herein, the method forproviding damping is shown, and the fact that the optimum value can bedetermined experimentally is emphasized.

Second Embodiment

FIG. 5 is a perspective view of a head suspension, showing a secondembodiment of a head suspension mechanism in accordance with the presentinvention. In the second embodiment, the aforementioned principle ofdynamic vibration absorber is applied so that second-order torsion modevibrations are damped efficiently. In FIG. 5 and the following figures,the same reference characters are applied to elements that aresubstantially the same as those in the related art, and the explanationthereof is omitted.

A head suspension 31 in accordance with the present invention has asuspension main frame 33 formed of a thin metal sheet.

In a base part of the suspension main frame 33, a leg portion 33 a isformed so as to provide flexural elasticity by blanking a central partthereof. Also, at both edges of the suspension main frame 33, flangeportions 33 b erecting along the edges are formed. Further, a pair ofdual tongue-shaped elastic plates 33 c formed into a cut-and-raise shapesymmetrically with respect to the center axis are formed as anadditional vibration system for playing a role of dynamic vibrationabsorber in locations close to the base part of the suspension mainframe 33. The root of the dual tongue-shaped elastic plate 33 c isintegrally connected to this suspension main frame 33. The location ofthis root corresponds to an antinode of second-order torsional vibrationmode. Therefore, the dual tongue-shaped elastic plate 33 c is vibratedmost effectively by an intended second-order torsion mode vibrations.Both of the length of the dual tongue-shaped elastic plate 33 c and thenatural frequency of flexural vibration are determined so as to meet thecondition coinciding with the second-order torsional vibration mode inaccord with the theory of dynamic vibration absorber.

The dual tongue-shaped elastic plate 33 c must also be provided with adamping effect. As a method for providing damping effect, there areavailable a method in which a damping material is bonded onto the uppersurface (or upper and lower surfaces) of the dual tongue-shaped elasticplate 33 c as in the method shown in FIG. 2 for the sway mode and amethod in which a part or the whole of a clearance between the dualtongue-shaped elastic plate 33 c and the suspension main frame 33 isfilled with a visco-elastic material as in the method shown in FIG. 4.Since the damping effect has the optimum value as described above, theoptimum value should be determined experimentally. The dualtongue-shaped elastic plate 33 c is formed by blanking the surroundingportion thereof, and the forming method is almost the same as theexisting cut-and-raise fabrication. In this case, however, the dualtongue-shaped elastic plate 33 c is normally located in the same planeas the suspension main frame 33. Therefore, the term “cut-and-raise” inthis case is used for convenience as a fabrication method, and actuallythe “raise” fabrication is not performed.

Third Embodiment

FIG. 6 is a perspective view of a head suspension, showing a thirdembodiment of a head suspension mechanism in accordance with the presentinvention. The third embodiment is configured so that the sway mode andsecond-order torsion mode vibrations are damped at the same time by thesame principle of dynamic vibration absorber.

In FIG. 6, in a base part of a suspension main frame 43, a leg portion43 a is formed so as to provide flexural elasticity by blanking acentral part thereof. Also, at both edges of the suspension main frame43, flange portions 43 b erecting along the edges are formed. Further, afree end portion 43 c that is made free toward the base part side bycutting substantially into an L shape is formed in a location close tothe tip end of the flange portion 43 b. Also, a pair of dualtongue-shaped elastic plates 43 d formed into a cut-and-raise shapesymmetrically with respect to the center axis are formed as anadditional vibration system for playing a role of dynamic vibrationabsorber in locations close to the base part of the suspension mainframe 43. The free end portion 43 c is substantially the same as thefree end portion 23 c, and the dual tongue-shaped elastic plate 43 d issubstantially the same as the dual tongue-shaped elastic plate 33 c.

Thereupon, the free end portion 43 c is the additional vibration systemfor playing a role of dynamic vibration absorber for damping the swaymode vibrations, and the dual tongue-shaped elastic plate 43 d is theadditional vibration system for playing a role of dynamic vibrationabsorber for damping the second-order torsion mode vibrations. These twostructures each are the same as the above-described additional vibrationsystems, and both of the length of vibration plate and the naturalfrequency of flexural vibration are determined so as to meet thecoincidence condition in accord with the theory of dynamic vibrationabsorber. In order to effectively damp the sway mode and second-ordertorsion mode vibrations at the same time, the additional vibrationsystem for damping the second-order torsion mode vibrations is firstdetermined, and then the additional vibration system for damping thesway mode vibrations is determined by the above-described method. Byrepeating this procedure, the final simultaneous coincidence conditionis determined.

Thus, for the head suspension mechanism with a dynamic vibrationabsorber, which is configured by the free end portion 23 c and thedamping material 23 d configured, for example, as shown in FIG. 1, thelength L2 of the beam part a of the free end portion 23 c is adjusted tomake design so that the natural frequency of the flexural vibration isin proximity to the natural frequency of sway mode vibration of the headsuspension mechanism, and the damping material 23 d is designed so thatthe damping effect is optimal. Thereby, the vibration energy of swaymode is absorbed by the damping material 23 d, and the resonanceamplitude multiplying factor (Q value) can be reduced. Hereunder, theeffect is explained by using a result of analysis of an actualsuspension mechanism performed by the finite element method.

FIG. 8 is a graph showing an amplitude of a compliance frequencyresponse function (FRF) of displacement amplitude in the Y and Zdirections. The amplitude was observed at a head position 9 a on aslider 9 shown in FIG. 7 at the time when a harmonic exciting force wasapplied in the Y and Z directions of point A of the suspension mainframe 23 in the head suspension mechanism with a dynamic vibrationabsorber for damping the sway mode resonance shown in FIG. 1. Thisamplitude is called an FRF amplitude.

In this analysis, since the slider 9 simulates an actual floating state,the Z direction is supported by elastic springs at pad positions 9 b, 9c and 9 d. A solid line in the graph of FIG. 8 indicates an FRF in thecase where the above-described coincidence condition is met and designis made so that damping is nearly optimal, and a broken line in thegraph indicates an FRF in the case where the dynamic vibration absorberis not provided. In this case, a length from the fixed plate 2 of thesuspension main frame 23 is 15.04 mm, and the thickness thereof is 38μm. The thickness of the free end portion 23 c of dynamic vibrationabsorber is 38 μm, the width of the beam part a is 0.3 mm, and thelength that meets the coincidence condition is 1.20 mm. Also, the widthof the connecting part is 0.3 mm. The optimum damping effect was 3×10⁻⁶in terms of equivalent tan δ value of Young's modulus of the beam part.

In the graph of FRF amplitude on the Y axis of FIG. 8, it is found thatthe highest-level frequency component P1 in the case where the dynamicvibration absorber is not provided, as indicated by the broken line,represents the sway mode, and the provision of the dynamic vibrationabsorber changes P1 to P2, decreasing the maximum value to 16 dB. Thesecond-order torsion mode is represented by frequencies indicated by P3and P4, and the FRF amplitude in the Y-axis direction is not changed bythe provision of dynamic vibration absorber. However, the FRF amplitudein the Z direction is larger than the amplitude in the case where thedynamic vibration absorber is not provided. In FIG. 8, P5 and P6represent the first-order bending mode, and P7 and P8 represent thethird-order bending mode. In both modes, the natural frequency isdecreased slightly by the provision of dynamic vibration absorber, butthe peak amplitude is decreased in both of the Y direction and the Zdirection.

As can be seen from the above description, in the head suspensionmechanism with the L-shaped dynamic vibration absorber, the vibrationamplitude in the Y direction in which a track shift occurs decreases,and the amplitude of sway mode, which poses the biggest problem, isdecreased significantly, and is kept at 20 dB or lower. Therefore, thehead suspension mechanism with the L-shaped dynamic vibration absorbershown in FIG. 1 has an advantage of being capable of damping sway modevibrations generated by wind disturbance and hence enabling high trackdensity recording.

FIG. 9 is a graph showing the FRF amplitude of the head suspensionmechanism with the dynamic vibration absorber, aiming at damping thesecond-order torsion mode vibrations, which is shown in FIG. 5. A solidline indicates the case where the dynamic vibration absorber isprovided, and a broken line indicates the case where the dynamicvibration absorber is not provided. In this case, the dimensions of thesuspension main frame are as described above, the width of the dualtongue-shaped elastic plate of dynamic vibration absorber is 0.2 mm, andthe length that meets the coincidence condition of dual tongue-shapedelastic plate is 1.35 mm. Also, the optimum damping effect was 3×10⁻⁶ interms of equivalent tan δ value of Young's modulus of the beam part.

As can be seen from these figures, the peak amplitude P3 of thesecond-order torsion mode vibration without the dynamic vibrationabsorber in the Z direction is decreased to an amplitude P4 by theprovision of the dynamic vibration absorber. This fact reveals that inthe case where the second-order torsion mode vibrations, especially thevibrations in the Z direction, which pose a problem, the resonanceamplitude can be decreased significantly by providing the dynamicvibration absorber having the construction as shown in FIG. 5.

FIG. 10 is a graph showing the FRF amplitude of the head suspensionmechanism provided with two dynamic vibration absorbers to damp the swaymode vibrations and the second-order torsion mode vibrations at the sametime, which is shown in FIG. 6. A solid line indicates the case wherethe dynamic vibration absorber is provided, and a broken line indicatesthe case where the dynamic vibration absorber is not provided. In thiscase, the dimensions of the suspension main frame are as describedabove. The thickness of the free end portion 23 c of the L-shapeddynamic vibration absorber aiming at damping the sway mode vibrations,the width of the beam part a, and the width of the connecting part werethe same as described above, and the length of the beam part a, whichmet the coincidence condition, was 1.23 mm. On the other hand, thelength of the dual tongue-shaped elastic plate 33 c of dynamic vibrationabsorber aiming at damping the second-order torsion mode vibrations,which met the coincidence condition, was 1.41 mm. Also, the optimumdamping effect was, as described above, 3×10⁻⁶ in terms of equivalenttan δ value of Young's modulus of the beam part.

From these figures, it is found that the peak amplitude P1 of the swaymode vibration and the peak amplitude P3 of the second-order torsionmode vibration in the case where no dynamic vibration absorber isprovided decrease to P2 and P4, respectively, by providing the dynamicvibration absorbers that meet the coincidence condition at the sametime. Also, it is found that the maximum value of FRF amplitude in the Ydirection of the sway mode vibration decreases to 14 dB. Further, it isfound that in the second-order torsion mode, not only the amplitude inthe Z direction can be decreased, but also the maximum value ofamplitude in the Y direction can be decreased to 15 dB as compared withthe case where the dynamic vibration absorber is provided to damp thesecond-order torsion mode vibrations only, as shown in FIG. 9.

As is apparent from the above description, the head suspension mechanismprovided with the plurality of dynamic vibration absorbers aiming atdamping both of the sway mode vibrations and the second-order torsionmode vibrations at the same time, which is shown in FIG. 6, decreasesthe sway mode vibrations that pose a problem of track shift and thevibrations in the Y direction due to the second-order torsion mode, andachieves a very great effect as compared with the construction providedwith the dynamic vibration absorber aiming at damping the single modevibrations. Therefore, it is preferable to design the head suspensionmechanism provided with the plurality of dynamic vibration absorbersaiming at damping vibrations of a plurality of modes as shown in FIG. 6.

As can be seen from the above description, the dynamic vibrationabsorber having the free end portion 23 c having the fixed part near thetip end of the head suspension mechanism has a characteristic ofefficiently performing the function of additional vibration system forrestraining the vibrations in the track shift direction of the headsuspension mechanism, so that an efficient dynamic vibration absorbereffect can be achieved by a little additional mass. Also, the free endportion 23 c has the same thickness as that of the suspension main frame23, and thus can be produced integrally from the same metal sheetmaterial, so that the manufacturing cost is low.

The paired dual tongue-shaped elastic plate having the fixed part at theantinode of torsion mode of the head suspension mechanism has acharacteristic of efficiently performing a function of additionalvibration system restraining vibrations in the out-of-plane direction oftorsion mode in the head suspension mechanism. Also, the additional massis little, and hence an efficient dynamic vibration absorber effect canbe achieved. The dual tongue-shaped elastic plate has a characteristicof being manufactured at a low cost because it can be prepared merely bycutting the suspension main frame partially into a U shape.

Furthermore, by combining the above-described two design plans, adynamic vibration absorber effect can be achieved which is greater thanthe sum of effects at the time of design aiming at damping a single modeonly. Actually, a design is effective in which a plurality of dynamicvibration absorbers aiming at damping vibrations of many modes areprovided.

The effects of the present invention have been explained regarding thehead suspension mechanism having no restraining multilayer visco-elasticmaterial, as shown in FIG. 2. However, the present invention canadditionally be applied to the head suspension mechanism with therestraining multilayer visco-elastic material, which is now being used,shown in FIG. 3. FIG. 11 shows another embodiment in which the free endportion 43 c and the dual tongue-shaped elastic plate 43 d are appliedto the head suspension mechanism with the restraining multilayervisco-elastic material shown in FIG. 6. In the embodiment shown in FIG.11, there is provided a damping plate 45 to restrain the sway mode andsecond-order torsion mode vibrations that are not enough restrained bythe restraining multilayer visco-elastic material only.

FIG. 11 shows an embodiment in which the restraining multilayervisco-elastic material is affixed to the embodiment of the presentinvention shown in FIG. 6. However, the restraining multilayervisco-elastic material can be affixed to the head suspension mechanismwith the L-shaped dynamic vibration absorber for damping the sway modevibrations in accordance with the present invention, which is shown inFIG. 1. Also, it is apparent that the restraining multilayervisco-elastic material can be affixed to the head suspension mechanismwith the dual tongue-shaped elastic plate shaped dynamic vibrationabsorber for damping the second-order torsion mode vibrations, which isshown in FIG. 6.

In the above-described embodiments, there has been disclosed the headsuspension mechanism having the flange portion 23 b extendingsubstantially over the whole length of the suspension main frame 23.However, the flange portion 23 b is not necessarily needed, and a flangeportion may be provided only a portion where the free end portion 23 cshown in FIG. 1(C) is substantially provided.

As described above, the suspension main frame 23, 33, 43 in accordancewith the present invention is provided with at least either one of theL-shaped dynamic vibration absorber and the dual tongue-shaped elasticplate shaped dynamic vibration absorber. Thereupon, the head suspensionmechanism can damp the sway mode and second-order torsion modevibrations that have not been enough damped conventionally by therestraining multilayer visco-elastic material. Therefore, the headsuspension mechanism achieves an effect of restraining the sway modevibrations resulting in a track shift caused by an air flow. Also, theL-shaped dynamic vibration absorber (the free end portion 23 c, 43 c)and the dual tongue-shaped elastic plate 33 c, 43 d in accordance withthe present invention have a characteristic of being capable of beingformed integrally of the same metal sheet as the suspension main frames23, 33, 43. Therefore, a head suspension mechanism that can bemanufactured at a low cost and is not subjected to wind disturbancevibrations can be provided.

1. A head suspension mechanism which supports a head slider rigidly inan in-plane direction and softly in an out-of-plane direction,comprising: a suspension main frame (23) having an elastic cantileverthin sheet extending horizontally to provide a load force to the headslider; a support portion (4) at a terminal end of the suspension mainframe (23) to support the head slider; a connecting portion at each ofthe two opposite edges of the suspension main frame (23) at the tip endportion, extending vertically from the opposite edges; and a free endportion (23 c) connected to each of the connecting portions, each freeend portion extending from an upper end of the connecting portion in ahorizontal direction opposite the tip end portion to form an L-shapewith a space between a lower edge of the free end portion and an edgesurface of the suspension main frame (23), wherein damping material isprovided on each free end portion, a dampening effect of the dampingmaterial and the length of the flange portion forming a sway modedynamic vibration absorber.